The goal of vibration isolation systems is to isolate a supporting base from the vibration of an isolated mass caused by a perturbing force, i.e., lowering the force transmitted to the base, while avoiding excessive vibrating motion of the isolated mass. Frequently the isolated mass (e.g., a motor) must also be isolated from the motion of the supporting base (e.g., the chassis of a vehicle) so that the effect of supporting base perturbation on the isolated, mass is minimized. Therefore, isolation systems attempt to lower the transmissibility of vibrations from the isolated mass to the supporting base and vice versa. Passive mounts/springs with negligible damping and low stiffness may be effective in isolating vibration at relatively higher frequencies but may have poor shock isolation characteristics (e.g., abrupt, discontinuous perturbations).
Isolation systems having mounting structures with negligible damping may have the lowest transmissibility at high, off-resonant frequencies. The excessive transmissibility at resonance in such underdamped systems is normally addressed by making their resonant frequencies well below the lowest vibration excitation frequency. Softer mounts (mounts with lower stiffness) may decrease both the transmission of force from the vibrating mass to the base and the transmission of shock inputs at the base to the mass over a reasonably large range in frequencies.
In many isolation applications the isolated mass and supporting base may be subject to both vibration, e.g., the force of operating an internal combustion engine, and shock excitation, e.g., the supporting base excitation in a shipboard machine caused by choppy waters. Shock excitation has normally a broadband spectrum, so some energy at the resonant frequency(ies) of the isolated system may be present. If the mounting structures of the isolation system are highly underdamped, the vibration amplitude at this (these) frequency(ies) may become excessive and cause damage to the isolated system. This problem may be even worse when the shock excitation is not random, but periodic/rhythmic with a harmonic at one of the resonant frequencies of the isolated mass. In this case, the isolation system may amplify the vibration instead of abating it.
Stiffness of the mounts of an isolation system is also a factor in isolation effectiveness, notably at low frequencies. The softer the mount, the higher is its low frequency vibration isolation performance. However, improved isolation using soft mounting structures is achieved at the expense of excessive low-frequency motion of the isolated machine in response to shock disturbances. For example, a diesel engine during start up and shut down experiences excessive undesirable motion due to shock excitation, thereby straining the plumbing, wiring, and exhaust connections to the engine.
Current passive isolation systems do not adequately provide isolation at both high frequency vibrations and shock excitations. For example, double (two-stage) mounting, while it may be effective at high frequencies, has less than desirable low-frequency isolation effectiveness. Additionally, double mounting may impose unfavorable weight and space penalties. Other isolation systems add tuned damping to the isolator at the resonant frequency of the isolated system, such as hydraulic engine mounts used in automobiles. Although applicable as an engine mount where the mass of the engine is not a time varying parameter, the use of passive tuned damping in applications such as semi-trailer truck cab isolation and seat suspension in agricultural and earth-moving machinery having time-varying mass is not very practical.
Accordingly, alternative isolation systems are desired.